Combustion engine waste heat powered air-conditioning system

ABSTRACT

An air conditioning system in which a high-pressure refrigerant vapor turbine is driving a low-pressure high-speed centrifugal compressor both supported on liquid refrigerant hydrostatic journal bearings. Due to required turbine miniaturization, the turbine blades surface finish and blade accuracy are of critical importance in order to produce high turbine adiabatic efficiency.

CROSS REFERENCE TO RELATED APPLICATIONS

This application claims the benefit of Provisional Patent Application Ser. No. 61/517,1-2 filed Apr. 13, 2011.

FIELD OF THE INVENTION

This invention relates to air conditioning systems and in particular to such systems based on the use of waste heat.

BACKGROUND OF THE INVENTION

Conventional air-conditioning systems in motor vehicles utilize predominantly direct driven refrigerant compressors which provides suction pressure causing evaporation of liquid refrigerant in evaporator that produces cooling capacity to the air flow circulated through the vehicle interior. Compressor vapor discharge is condensed in a condenser where it is usually cooled by ambient air. Condensed liquid is further expanded through a throttle valve back into the evaporator thus forming a closed loop.

Typical direct driven air-conditioning system in a hot climate uses approximately 1 kW of engine power per 1 ton cooling capacity, equivalent to 12,000 BTU/hr. Current retail price of gasoline and diesel fuels in US is approximately 5 dollars per gallon or about 85 cents per pound of fuel. Typical 4 ton automotive air-conditioning requires approximately 4 kW or 5.4 HP worth of engine power. With average specific fuel consumption of approximately 0.55 lb/HP-hr this translates into approximately three dollars per hour fuel cost just to run the vehicle air conditioning. Besides the fuel cost this contributes to vehicle exhaust emission and needless use of world fuel supply.

Applicant has developed over the past decade a miniature very high speed, high efficiency turbine technology such as shown in the Applicant's U.S. Pat. No. 5,924,286. This proven technology is directly applicable to the waste heat powered refrigeration system, subject of this invention.

Therefore, there is a great need for efficient and low cost, engine waste heat powered air-conditioning system.

SUMMARY OF THE INVENTION

Present invention provides an air conditioning system in which a high-pressure refrigerant vapor turbine is driving a low-pressure high-speed centrifugal compressor both supported on liquid refrigerant hydrostatic journal bearings. Due to required turbine miniaturization, the turbine blades surface finish and blade accuracy are of critical importance in order to produce high turbine adiabatic efficiency. Because of the possibility of occasional refrigerant liquid mist passing through turbine blades it is extremely important that turbine blades are resistant to liquid droplets erosion. Turbine efficiencies of 80% and higher have been achieved with very high speed turbines similar in size to the refrigerant turbine subject of this invention. Turbine blades are manufactured of a Du Pont high-temperature Vespel plastic encased in metal wheel for dimensional and thermal stability. This technology, including specific turbine design details is shown in Applicant's U.S. Pat. No. 5,924,286. Those blades were proven as extremely erosion resistant over years of operation in hydraulic fluids with up to 600 ft/sec fluid velocities.

The system is powered by waste heat of a combustion engine having a coolant pump, a radiator and a coolant flow control valve said air conditioning system. The system includes a refrigerant boiler in fluid communication with hot coolant exiting said combustion engine, a hot coolant flow control valve adapted to control flow of the hot coolant to the refrigerant boiler, a coolant return line for returning fluid exiting the refrigerant boiler to an inlet of the coolant pump, a high-pressure refrigerant pump adapted to provide the refrigerant boiler with high-pressure refrigerant, a refrigerant compressor defining a low-pressure refrigerant vapor inlet and a high-pressure fluid refrigerant outlet, a first refrigerant vapor turbine, in vapor communication with the refrigerant boiler, driven by high-pressure refrigerant vapor produced by the refrigerant boiler and adapted to drive the refrigerant compressor, a refrigerant condenser adapted to condense refrigerant vapor discharged by the first refrigerant vapor turbine and the refrigerant compressor, and a refrigerant evaporator in high-pressure refrigerant fluid communication with the refrigerant compressor outlet and in low pressure refrigerant vapor communication with the low-pressure refrigerant compressor inlet and adapted to provide cooling of an enclosed space as a consequence of evaporation of evaporation to the high-pressure refrigerant.

In preferred embodiments, the engine waste heat contained in the engine cooling water loop is transmitting the heat to high pressure refrigerant boiler generating the vapor flow that drives the turbine. As shown in tabulated data bellow, the engine coolant heat input to the refrigerant boiler is more than sufficient to power the novel air conditioning system down to idle speeds in most vehicular applications. Optionally, the system can be augmented with the engine exhaust heat (not shown) if necessary. In a typical engine about ⅓ of heat input into the engine is rejected via engine coolant and ⅓ is contained in the engine exhaust.

In some embodiment the costs associated with electric driven pump or belt driven pump systems can be eliminated, saving an additional 0.5 HP of engine power and greatly improves reliability of the overall system by simplicity of zero leak self-contained bearing system in the FIG. 7 turbo-pump. Mass production cost of the FIG. 7 turbo-pump is estimated to be comparable or better than other commercial electric or belt driven refrigerant pumps.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows overall system diagram that includes engine and its cooling system connected to refrigerant boiler, turbine-compressor, evaporator, condenser and control valves.

FIG. 2 is a cross sectional drawing of the turbo-compressor.

FIG. 3A is a thermodynamic diagram of the refrigerant loops.

FIG. 3B compares radiator flow to A/C boiler flow.

FIGS. 4A and 4B are cross sectional drawings of a typical plastic-metal turbine wheel design as shown in U.S. Pat. No. 5,924,286.

FIGS. 5A and 5B show design details of a typical 80 percent plus efficient turbine nozzle as shown in U.S. Pat. No. 5,924,286.

FIGS. 6A and 6B show design details of a typical 80 percent plus efficient turbine blade as shown in U.S. Pat. No. 5,924,286.

FIG. 7 is a drawing showing a second preferred embodiment of the present invention.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENTS

Preferred embodiments of the present invention are described by reference to the drawings.

FIG. 1 shows engine radiator 55 receiving hot engine coolant from engine 51 via hot coolant pipe 61 and radiator flow control valve 62. Pump 52 forces coolant through engine 51 radiator control valve 62 and radiator 55. A pressure drop across valve 62 forces hot engine coolant into refrigerant boiler 31 via line 59 and hot coolant control valve 53. Line 57 provides for engine coolant return from the refrigerant boiler 31 to the radiator return line 58 and to inlet of engine coolant pump 52. Hot coolant control valve 53 adjusts the flow rate of the hot coolant into the refrigeration boiler 31 as air conditioning load requires. Radiator flow control valve 62 reduces coolant flow to radiator 55 as the engine load and speed decreases while hot coolant control valve 53 opens up more as air conditioning load requires. At maximum air conditioning load heat rejection into refrigerant boiler 31 stays constant while amount of heat rejected from radiator 55 changes as function of engine power level as shown in Table 1.

Water heated in engine 51 is circulated via line 59 and hot coolant control valve 53 into refrigerant boiler 31 where high pressure refrigerant received from electric driven or engine shaft driven refrigerant pump 32 via line 33 is boiled off. Refrigerant vapor generated in refrigerant boiler 31 flows via line 28 and further on via turbine control valve 27 into refrigerant turbine 13 which is directly driving refrigerant compressor 12. Refrigerant vapor discharged out of the refrigerant turbine 13 joins the refrigerant vapor discharged by the refrigerant compressor 12 via line 15 into the line 19. Total refrigerant flow discharged by the refrigerant turbine 13 and by the refrigerant compressor 12 flows via line 21 into the refrigerant condenser 22 which is cooled by ambient air or air-water mixture or cooling water or swimming pool water. Refrigerant line 23 provides liquid refrigerant flow into line 26 and further on via refrigerant cooler control valve 18 into the refrigerant evaporator 17 which provides cooling to the air conditioned air by means of refrigerant evaporation in the refrigerant evaporator 17. Refrigerant vapor flow is than compressed by the refrigerant compressor 12 into the line 15 and further on as previously described.

Performance and Design Details of the Preferred System Embodiment Engine Coolant Capacity Driving the Refrigerant Loop

Thermal analysis of engine coolant and refrigerant system preferred embodiment was conducted for constant air conditioning load of 4 ton cooling capacity and variable engine power levels for typical 250 HP and 500 HP heavy duty truck diesel engines.

Tables 1 and 2 below show the effect of engine part load on percentage of total engine coolant used by the refrigeration boiler to drive refrigerant turbine-compressor producing 4 ton cooling capacity. Hot coolant temperature is conservatively assumed to be 220 deg. F. which is equivalent to 4 psig engine coolant pressure.

TABLE 1 250 HP Engine Engine load % 100 75 50 25 5 Engine power HP 250 187.5 125 62.5 12.5 Estimated engine SFC 0.45 0.45 0.48 0.65 1.30 (lb/HP-hr) Fuel flow lb/hr 112.5 84.4 60 40.6 16.25 Total fuel heat BTU/hr 1.97 1.47 1.05 0.75 0.28 (×10⁻⁶) Cooling water heat 0.59 0.44 0.31 0.21 0.085 BTU/hr (×10⁻⁶) Refrigerant boiler 0.077 0.077 0.077 0.077 0.077 BTU/hr (×10⁻⁶) Engine radiator BTU/hr 0.513 0.363 0.233 0.133 0.008 (×10⁻⁶)

TABLE 2 500 HP Engine Engine load % 100 75 50 25 5 Engine power HP 500 375 250 125 25 Estimated engine SFC 0.45 0.45 0.48 0.65 1.3 (lb/HP-hr) Fuel flow lb/hr 225 168 120 81 32 Total fuel heat BTU/hr 3.94 2.95 2.10 1.42 0.57 (×10⁻⁶) Cooling water heat 1.18 0.89 0.63 0.43 0.17 BTU/hr (×10⁻⁶) Refrigerant boiler 0.077 0.077 0.077 0.077 0.077 BTU/hr (×10⁻⁶) Engine radiator BTU/hr 1.10 0.81 0.55 0.35 0.093 (×10⁻⁶)

Above analysis shows that relatively small amount of engine waste heat is required to drive 4 ton capacity preferred embodiment air conditioning system.

Thermal analysis of the preferred system embodiment was conducted using Du Pont's refrigerant HCFC-124 which for ecological reasons replaces old R-114 refrigerant. Refrigerant compressor efficiency of 75% and refrigerant turbine efficiency of 80% were assumed in the system analysis.

Table 3. below show the effect of air-conditioning cooling capacity on thermal rating of the refrigerant evaporator 17, refrigerant boiler 31 and on the optimum size and speed of the refrigerant turbine 13 and refrigerant compressor 12.

As shown in Table 3, the optimum size of turbine and compressor wheels is quite small and operating RPM high requiring advanced high speed miniature turbine technology.

TABLE 3 COOLING CAPACITY (ton) 2 4 10 100 Refrigerant 24,000 48,000 120,000 1,200,000 evaporator heat load (BTU/hr) Refrigerant boiler 35,530 71,600 177,651 1,776,515 heat load (BTU/hr) Refrigerant turbine- 1.47 2.94 7.35 73.54 compressor power (kW) Refrigerant turbine- 99,929 70,661 44,690 14,132 compressor (RPM) Refrigerant turbine 15.67 22.16 35.04 110.81 wheel diameter (mm) Refrigerant 36.01 50.93 80.53 254.66 compressor diameter (mm)

FIG. 3A shows typical pressure-enthalpy (P-H) refrigerant process diagram and how it relates to the system component numbers shown in FIG. 1. Line 17 shows the amount of heat absorbed per one pound of refrigerant in the refrigerant evaporator 17. Line 12 shows compression process in the refrigerant compressor 12. Line 22 shows the amount of heat rejected per one pound of refrigerant in the refrigerant condenser 22. Line 18 represents throttling process in the refrigerant cooler control valve 18. Line 32 represents pumping process in the refrigerant pump 32. Line 31 shows the amount of heat absorbed per one pound of refrigerant in the refrigerant boiler 31. Line 13 shows expansion work process in the refrigerant turbine 13.

FIG. 3B shows the estimated coolant flow split between the radiator flow and the refrigerant flow.

Preferred embodiments of the present invention utilize the same turbine design as is described in Applicant's U.S. Pat. No. 5,928,286 which describes a hydraulic supercharger system and is incorporated herein by reference. FIGS. 4A and 4B are cross sectional drawings of a typical plastic-metal turbine wheel design as shown in U.S. Pat. No. 5,924,286. FIGS. 5A and 5B show design details of a typical 80 percent plus efficient turbine nozzle as shown in U.S. Pat. No. 5,924,286. FIGS. 6A and 6B show design details of a typical 80 percent plus efficient turbine blade as shown in U.S. Pat. No. 5,924,286.

Second Preferred Embodiment

FIG. 7 shows second preferred system embodiment which is similar to the first preferred embodiment shown in FIG. 1 with exception that refrigerant pump 32 which is electric driven or belt driven by the engine shown in FIG. 1 is being replaced by high speed refrigerant pump 81 that is driven by refrigerant vapor turbine 82 as shown in FIG. 7. Refrigerant pump 81 and refrigerant vapor turbine 82 rotor is supported on liquid refrigerant hydrostatic journal bearings in the same fashion as refrigerant turbo-compressor 12 rotor shown in FIG. 3, thus avoiding need for oil lubrication.

Thermal analysis of engine coolant and refrigerant system has shown that amount of engine coolant waste heat available in typical heavy duty diesel engine is more than sufficient to generate refrigerant vapor in refrigerant boiler 31 to drive both compressor drive turbine 13 and refrigerant pump drive turbine 82.

In case of 4 ton air conditioning system the compressor drive turbine 13 produces 5.5 HP@70,600 rpm and the boiler feed pump drive turbine 82 produces only 0.5 HP@37,000 rpm thus requiring approximately 10% refrigerant vapor flow of the compressor drive turbine 13 flow.

Table 4 below shows optimized parameters of the refrigerant pump 81 driven by refrigerant vapor turbine 82 for a 4 ton HCFC-124 air-conditioning system.

TABLE 4 Pump inlet pressure (psia) 110 Pump pressure rise (psid) 190 Pump flow rate (gpm) 2.51 Pump/Turbine power (HP) 0.51 Pump speed (rpm) 37,790 Pump impeller diameter (mm) 19.16 Turbine wheel diameter (mm) 40.8

Start-up of system shown in FIG. 7 requires initiation of refrigerant vapor flow through refrigerant vapor turbine 82. To achieve this, the engine must be warmed-up to its operating temperature with refrigerant boiler 31 filled with refrigerant fluid. At start-up the refrigerant fluid is contained in the refrigerant boiler 31 by closing turbine control valve 27 and closed check valve 74 installed in the pump discharge line 75 preventing reverse flow through pump discharge line 75. With hot coolant control valve 53 partially open, the heat input into refrigerant boiler 31 starts to generate refrigerant vapor flow which via refrigerant boiler discharge line 28 and with closed turbine control valve 27 forces the refrigerant vapor flow through the refrigerant vapor turbine 82 starting rotation of refrigerant pump 81 and pressurizing the refrigerant pump discharge line 75 which opens the check valve 74 and starts refrigerant circulation through refrigerant boiler 31. Further opening of the hot coolant control valve 53 increases the amount of refrigerant vapor flow generated by refrigerant boiler 31 at which point the turbine control valve 27 starts opening and driving the refrigerant vapor turbine 13 which drives the refrigerant compressor 12 and starts producing desired amount of air-conditioning cooling.

Variations

The present invention has been described above in terms of preferred embodiments. Persons skilled in the air condition and motor vehicles arts will understand that many changes and additions could be made within the general scope of the present invention. Applicant expects that the big market for systems according to the present invention will be motor vehicles, but the invention can be adapted for utilization of other sources of waste heat or even heat sources that are not waste heat, such as a solar concentrator. For solar powered air conditioning an electric pump replacing pump 52 in FIG. 1 or FIG. 7 could be used to circulate coolant through the concentrator directly to and through refrigerant boiler 31 back to the input of the electric pump. An alternative would be to replace the electric pump with a high speed refrigerant pump such as refrigerant pump 81 shown in FIG. 7. In this case the system could be completely solar powered. Electric controls could be powered with a solar panel or a thermoelectric device.

Therefore the scope of the present invention should be determined by the appended claims and their legal equivalence and not by the specific embodiment described above. 

What is claimed is:
 1. An air conditioning system powered by waste heat of a combustion engine having a coolant pump, a radiator and a coolant flow control valve said air conditioning system comprising: A) refrigerant boiler in fluid communication with hot coolant exiting said combustion engine, B) a hot coolant flow control valve adapted to control flow of the hot coolant to the refrigerant boiler, C) a coolant return line for returning fluid exiting the refrigerant boiler to an inlet of the coolant pump, D) a high-pressure refrigerant pump adapted to provide the refrigerant boiler with high-pressure refrigerant, E) a low-pressure high-speed refrigerant compressor defining a low-pressure refrigerant vapor inlet and a high-pressure fluid refrigerant outlet, F) a first high-pressure refrigerant vapor turbine, in vapor communication with the refrigerant boiler, driven by high-pressure refrigerant vapor produced by the refrigerant boiler and adapted to drive the refrigerant compressor, G) a refrigerant condenser adapted to condense refrigerant vapor discharged by the first refrigerant vapor turbine and the refrigerant compressor, and H) a refrigerant evaporator in high-pressure refrigerant fluid communication with the refrigerant compressor outlet and in low pressure refrigerant vapor communication with the low-pressure refrigerant compressor inlet and adapted to provide cooling of an enclosed space as a consequence of evaporation of evaporation to the high-pressure refrigerant, wherein the high-pressure refrigerant vapor turbine and the low-pressure high-speed centrifugal compressor are both supported on liquid refrigerant hydrostatic journal bearings.
 2. The system as in claim 1 wherein the high-pressure refrigerant pump is and electric powered pump.
 3. The system as in claim 1 wherein the high-pressure refrigerant pump is an engine shaft powered pump.
 4. The system as in claim 1 and also comprising a second refrigerant vapor turbine adapted to drive the high-pressure refrigerant pump.
 5. The system as in claim 4 wherein the second refrigerant vapor turbine comprises a rotor supported on liquid refrigerant hydrostatic journal bearings.
 6. The system as in claim 1 wherein the first refrigerant vapor turbine comprises a rotor supported on liquid refrigerant hydrostatic journal bearings.
 7. The system as in claim 1 wherein the system is adapted to provide about 4 tons of air conditioning.
 8. The system as in claim 1 wherein the first refrigerant vapor turbine is designed in accordance with the teachings of U.S. Pat. No. 5,924,286. 